U.S. patent application number 12/302486 was filed with the patent office on 2010-08-12 for continuously variable transmission.
Invention is credited to Robert Charles Downs.
Application Number | 20100199805 12/302486 |
Document ID | / |
Family ID | 38895068 |
Filed Date | 2010-08-12 |
United States Patent
Application |
20100199805 |
Kind Code |
A1 |
Downs; Robert Charles |
August 12, 2010 |
CONTINUOUSLY VARIABLE TRANSMISSION
Abstract
This invention provides continuously variable transmissions and
associated powertrains and automotive transmission systems.
Included among the transmission systems provided are efficient
transmissions for employing regenerative braking.
Inventors: |
Downs; Robert Charles; (La
Jolla, CA) |
Correspondence
Address: |
TIMOTHY L. SMITH
14740 CAMINITO PORTA DELGADA
DEL MAR
CA
92014
US
|
Family ID: |
38895068 |
Appl. No.: |
12/302486 |
Filed: |
May 30, 2007 |
PCT Filed: |
May 30, 2007 |
PCT NO: |
PCT/US2007/013003 |
371 Date: |
November 25, 2008 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
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60803386 |
May 30, 2006 |
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Current U.S.
Class: |
74/837 |
Current CPC
Class: |
Y10T 74/1675 20150115;
F16H 21/36 20130101; F16H 29/08 20130101; F16H 37/14 20130101 |
Class at
Publication: |
74/837 |
International
Class: |
F16H 21/20 20060101
F16H021/20 |
Claims
1. A continuously variable transmission comprising: an input shaft
that rotates relative to a reference frame; an input disk
comprising a face and a rotational axis at a center of the face,
wherein the input disk is operably connected to the input shaft
such that the input disk rotates about the rotational axis,
relative to the reference frame, when the input shaft rotates; a
crank pin movably attached to the face of the input disk, wherein
the crank pin is movable between at least a first position and a
second position on the face of the input disk; and an output
element that comprises: a bell crank that comprises an arm and a
pivot end; a connecting rod that connects the crank pin to an
attachment point on the arm of the bell crank; a one-way clutch
attached to the pivot end of the bell crank; and a clutch-driven
output shaft operably connected to the one-way clutch.
2. The continuously variable transmission of claim 1, wherein the
first position is at the rotational axis of the input disk.
3. The continuously variable transmission of claim 2, wherein the
maximum distance between the first position and the second position
is less than or equal to about eighty percent of the distance
between the pivot end and the attachment point of the bell
crank.
4. The continuously variable transmission of claim 1, wherein when
the crank pin is positioned at the rotational axis of the input
disk, rotation of the input shaft does not cause the connecting rod
to move.
5. The continuously variable transmission of claim 1, wherein when
the crank pin is positioned at a position other than the center of
the face of the input disk, rotation of the input disk causes the
connecting rod to move in a reciprocating substantially linear
motion, thereby causing the arm of the bell crank to oscillate in a
first direction and a second direction opposite to the first
direction, wherein movement of the arm of the bell crank in the
first direction is transmitted by the one-way clutch to the
clutch-driven output shaft and movement of the arm of the bell
crank in the second direction causes the one-way clutch to slip
such that the movement of the bell crank is not transmitted to the
clutch-driven output shaft.
6. The continuously variable transmission of claim 1, wherein the
transmission does not include a movable fulcrum between the input
shaft and the clutch-driven output shaft.
7. The continuously variable transmission of claim 1, wherein the
transmission does not include a lever between the input shaft and
the clutch-driven output shaft.
8. The continuously variable transmission of claim 1, further
comprising a driveshaft operably connected to the clutch-driven
output shaft, wherein rotational motion of the clutch-driven output
shaft is transmitted to the driveshaft.
9. The continuously variable transmission of claim 8, wherein the
rotational motion of the clutch-driven output shaft is transmitted
to the driveshaft by a rotation transmitting mechanism selected
from a gearset, a sprocket and chain, and a belt and pulley.
10. The continuously variable transmission of claim 1, wherein the
transmission comprises two or more output elements.
11. The continuously variable transmission of claim 10, wherein the
transmission comprises two output elements, each of which comprises
a connecting rod that is attached to the crank pin, wherein the
output elements are positioned approximately equidistant from each
other about the rotational axis of the input disk.
12. The continuously variable transmission of claim 10, wherein the
transmission comprises four output elements, each of which
comprises a connecting rod that is attached to the crank pin,
wherein the output elements are spaced approximately equidistant
from each other about the rotational axis of the input disk.
13. The continuously variable transmission of claim 10, wherein the
clutch-driven output shafts of each of the output elements are
operably connected to a driveshaft.
14. The continuously variable transmission of claim 13, wherein the
clutch-driven output shafts are operably connected to the
driveshaft by a gear mechanism.
15. The continuously variable transmission of claim 13, wherein the
clutch-driven output shafts are operably connected to the
driveshaft by a chain and sprocket mechanism.
16. The continuously variable transmission of claim 1, wherein the
input shaft comprises a crankshaft.
17. The continuously variable transmission of claim 1, further
comprising: at least a second bell crank that comprises an arm; at
least a second connecting rod that connects the crank pin to the
arm of the second bell crank; and at least a second one-way clutch
attached to the second bell crank; wherein the second one-way
clutch is operably connected to a clutch-driven shaft.
18. The continuously variable transmission of claim 17, wherein the
second one-way clutch is operably connected to the same
clutch-driven output shaft as the first one-way clutch.
19. The continuously variable transmission of claim 17, wherein the
second one-way clutch is operably connected to a second
clutch-driven output shaft.
20. The continuously variable transmission of claim 17, wherein the
connecting rod comprises a Scottish yoke that comprises a slotted
opening through which the crank pin is attached, and two ends, each
of which is attached to an arm of a bell crank.
21. The continuously variable transmission of claim 1, wherein the
transmission further comprises a backing plate to which the
clutch-driven output shaft is rotatably attached.
22. The continuously variable transmission of claim 21, wherein the
input shaft is fixed and the backing plate rotates about the input
shaft when a rotational force is applied to the clutch-driven
output shaft.
23. The continuously variable transmission of claim 22, wherein
when the crank pin is positioned at the center of the face of the
input disk the backing plate rotates at the same rotational speed
as the clutch-driven output shaft.
24. The continuously variable transmission of claim 22, wherein
when the crank pin is displaced from the center of the face of the
input disk the backing plate rotates at a reduced rotational speed
compared to the rotational speed of the clutch-driven output
shaft.
25. The continuously variable transmission of claim 22, wherein
when the crank pin is positioned on the face of the input disk a
distance from the rotational axis of the input disk that is greater
than fifty percent of the distance between the pivot end and the
attachment point of the bell crank and a rotational force is
applied to the backing plate, the clutch-driven output shaft
rotates in a direction opposite to that of the rotational
force.
26. The continuously variable transmission of claim 25, further
comprising an overrun/reversing mechanism interposed between the
clutch-driven outputshaft and a driveshaft, wherein the
overrun/reversing mechanism converts the rotational motion of the
clutch-driven output shaft to a rotational motion in an opposite
direction.
27. A regenerative braking system that comprises a driveshaft, a
flywheel, an engine, and: a) a first continuously variable
transmission configured to operate in overdrive mode to transmit
rotational motion from the driveshaft to the flywheel; b) a second
continuously variable transmission configured to operate in an
underdrive mode to transmit rotational motion from the engine or
the flywheel to a driveshaft; c) an engine input clutch; d) a
flywheel input clutch; and e) an output clutch.
Description
BACKGROUND OF THE INVENTION
[0001] A 1200 kg vehicle is typically designed with an engine with
a power output of approximately 100 Horsepower (HP). This is the
case in spite of the fact that only about 10 HP is needed to propel
the vehicle at a steady speed of 100 km/hr. The extra reserve power
is needed to accelerate the vehicle from rest and to propel the
vehicle up inclines.
[0002] This state of affairs is quite inefficient with regard to
fuel consumption as internal combustion engines achieve maximum
fuel efficiency when operating at or near their maximum power
output. A 100 HP engine is very inefficient while producing only 10
HP.
[0003] Hybrid electric vehicles offer improved efficiency by
reducing the size of the internal combustion engine (to about 70 HP
in this example) and supplementing the peak power requirements with
and electric motor/generator and battery system. (30 HP) The gas
engine is shut off completely during idling and very low speed
operation--consuming no fuel whatsoever. When maximum acceleration
is required, both the gas engine and electric motor/generator are
used together to obtain 100 HP. During most other operating
situations, the internal combustion engine operates alone and also
recharges the battery system by driving the motor/generator. In
these situations, the internal combustion engine operates closer to
its maximum efficiency point than a 100 HP engine would.
[0004] These two fuel saving modes represent a fairly significant
overall fuel saving over a conventional vehicle particularly if a
large portion of low speed and/or stop and go driving is being
performed.
[0005] Using a process known as regenerative braking can also save
some energy. In this process, the electric generator on the vehicle
is used to slow the vehicle during braking maneuvers, storing some
energy in the battery system that otherwise would have been
entirely wasted as heat in the friction brakes.
[0006] If it were possible to recover most or all of this vehicle
kinetic energy normally lost in braking, very large energy savings
would result. In practice, hybrid electric vehicles recover only a
fraction of this energy. This is primarily due to two factors.
First, the many steps of energy conversation each introduce
cumulative losses that add up to a significant loss. Kinetic
(mechanical) energy must first be converted to electric energy.
Then the electric energy must be transformed in the power
electronics. Then this transformed electric energy must be
converted to chemical energy in the battery system. Then to get
useful work from this stored energy, all these steps with their
accompanying losses must be reversed and repeated. If each step had
and average efficiency of 90%, the net efficiency of all of these
six steps would be (0.9).sup.6=50%.
[0007] Secondly, most practical battery systems have limitations on
the rate at which they can accept charging currents. A 1200 kg
vehicle traveling at 100 km/hr represents about 1,000,000 joules of
kinetic energy. To decelerate this vehicle to zero in a reasonable
10 seconds and store all of the energy means the battery system
would have to accept a power flux of 100 kilowatts for ten seconds.
This is about an order of magnitude beyond what a battery sized for
a hybrid vehicle could accept through charging.
[0008] If an energy storage system could be designed that would
capture and use most or all of the kinetic energy normally lost
during braking very large efficiencies could be realized. In
theory, a vehicle could be constructed with an internal combustion
engine downsized all the way to 10 or 20 HP. One key innovation
required to bring this concept to fruition is a continuously
variable transmission (CVT) with a very wide ratio coverage range
and mechanical efficiencies comparable to a geared transmission.
The invention described herein fulfils this and other needs.
SUMMARY OF THE INVENTION
[0009] The invention provides a continuously variable transmission
that includes: a) an input shaft that rotates relative to a
reference frame; b) an input disk comprising a face and a
rotational axis at a center of the face, wherein the input disk is
operably connected to the input shaft such that the input disk
rotates about the rotational axis, relative to the reference frame,
when the input shaft rotates; c) a crank pin movably attached to
the face of the input disk, wherein the crank pin is movable
between at least a first position and a second position on the face
of the input disk; and d) an output element that comprises: i) a
bell crank that comprises an arm and a pivot end, ii) a connecting
rod that connects the crank pin to an attachment point on the arm
of the bell crank, iii) a one-way clutch attached to the pivot end
of the bell crank, and iv) a clutch-driven shaft operably connected
to the one-way clutch.
[0010] In some embodiments, when the crank pin is positioned on the
face of the input disk at a position other than the center,
rotation of the input disk causes the connecting rod to move in a
reciprocating substantially linear motion, thereby causing the arm
of the bell crank to oscillate in a first direction and a second
direction opposite to the first direction, wherein movement of the
arm of the bell crank in the first direction is transmitted by the
one-way clutch to the clutch-driven shaft and movement of the arm
of the bell crank in the second direction causes the one-way clutch
to slip such that the movement of the bell crank is not transmitted
to the clutch-driven shaft.
[0011] The continuously variable transmissions can also include an
output shaft operably connected to the clutch-driven shaft, wherein
rotational motion of the clutch-driven shaft is transmitted to the
output shaft.
[0012] The transmissions can have one or more output elements
attached to the crank pin. Typically, when two or more output
elements are used, the output elements are positioned approximately
equidistant from each other about the rotational axis of the input
shaft. In multiple output element embodiments, the clutch-driven
shafts are generally connected to a common output shaft by, for
example, a gear mechanism, a chain and sprocket, or similar
means.
[0013] In some embodiments, the continuously variable transmission
include at least a second bell crank that comprises an arm, at
least a second connecting rod that connects the crank pin to the
arm of the second bell crank; and at least a second one-way clutch
attached to the second bell crank. The second one-way clutch is
operably connected to a clutch-driven shaft, which can be either
the same clutch-driven shaft to which is attached the first bell
crank or can be a second clutch-driven shaft.
[0014] The reference frames of the continuously variable
transmissions can include a backing plate to which the
clutch-driven shaft is attached. The input shaft in these
embodiments can be fixed, while the backing plate rotates about the
input shaft when a rotational force is applied to the clutch-driven
shaft. If the crank pin is positioned at the center of the face of
the input disk, the backing plate rotates at the same rotational
speed as the clutch-driven shaft. However, when the crank pin is
moved to a position other than the center of the input disk face,
the backing plate will rotate at a reduced rotational speed
compared to the rotational speed of the clutch-driven shaft.
[0015] The invention also provides a regenerative braking system
that includes a driveshaft, a flywheel, an engine, and: a) a first
continuously variable transmission configured to operate in
overdrive mode to transmit rotational motion from the driveshaft to
the flywheel; b) a second continuously variable transmission
configured to operate in an underdrive mode to transmit rotational
motion from the engine or the flywheel to a driveshaft; c) an
engine input clutch; d) a flywheel input clutch; and e) an output
clutch.
DESCRIPTION OF THE DRAWINGS
[0016] FIGS. 1A and B show a continuously variable transmission
that has a rotating input crankshaft 10 that is operably connected
to an input disk 20 such that the input disk rotates about its
center when the input crankshaft rotates. A crank pin 30 is movably
positioned on a face of the input disk. In FIG. 1A, the crank pin
is positioned at the axis of rotation of the input disk. In FIG.
1B, the crank pin is displaced to a second position on the face of
the input disk.
[0017] FIGS. 2A-B shows a view of a connecting rod 40 attached to
the crank pin 30. In FIG. 2A, the crank pin is positioned at the
axis of rotation of the input disk 20, so the connecting rod does
not move when the input disk rotates. In FIG. 2B, the crank pin is
displaced to a position away from the axis of rotation of the input
disk. Rotation of the input disk results in a reciprocating
movement of the opposite end of the connecting rod.
[0018] FIGS. 3A-E show a continuously variable transmission that
includes a bell crank 50 attached to an end of the connecting rod
40 opposite to the end that is attached to the crank pin 30. A
one-way clutch 60 is attached to the bell crank. In FIG. 3A, the
crank pin is positioned at the axis of rotation of the input disk
so rotation of the input disk does not result in any movement of
the bell crank and the output shaft 70 remains stationary. In FIGS.
3B-E, the crank pin has been repositioned to a second position on
the face of the input disk. Rotation of the input disk results in
reciprocating movement of the connecting rod, which in turn causes
a back and forth rotational movement of the bell crank 50. The
one-way clutch transmits only clockwise (in the illustrated
embodiment) rotation of the bell crank to the output shaft 70.
[0019] FIG. 4 shows a time versus rotational speed output diagram
of the continuously variable transmission shown in FIG. 3.
[0020] FIG. 5 shows an example of a continuously variable
transmission having two output elements, each of which includes a
connecting rod 40, a bell crank 50, a one-way clutch 60, and an
output shaft 70.
[0021] FIG. 6 shows a time versus rotational speed output diagram
of the continuously variable transmission shown in FIG. 5.
[0022] FIG. 7 shows an example of a continuously variable
transmission having four output elements.
[0023] FIG. 8 shows a time versus rotational speed output diagram
of the continuously variable transmission shown in FIG. 7.
[0024] FIG. 9 shows an example of a continuously variable
transmission having eight output elements.
[0025] FIG. 10 shows a time versus rotational speed output diagram
of the continuously variable transmission shown in FIG. 9.
[0026] FIG. 11 shows an example of a continuously variable
transmission having two output elements that both drive the same
output shaft 70.
[0027] FIG. 12 shows an example of a continuously variable
transmission having four output elements, two of which drive each
of two output shafts 70.
[0028] FIG. 13 shows an example of a continuously variable
transmission in which a scotch yoke 90 is caused to reciprocate by
the crank pin 30 when the input disk rotates. The scotch yoke
includes a slot 80 into which the crank pin is inserted.
[0029] FIG. 14 shows an example of a continuously variable
transmission that employs two scotch yokes to drive four output
shafts.
[0030] FIG. 15 shows another embodiment of a continuously variable
transmission that employs a scotch yoke to drive two output
shafts.
[0031] FIG. 16 shows an embodiment of a continuously variable
transmission that employs a scotch yoke to drive two output
shafts.
[0032] FIG. 17 shows another embodiment of a continuously variable
transmission that employs a scotch yoke to drive two output
shafts.
[0033] FIG. 18 shows a further example of a continuously variable
transmission that uses a scotch yoke to drive two output
shafts.
[0034] FIG. 19 shows one embodiment of a continuously variable
transmission acting in underdrive mode. A rotational input of 1000
rpm is applied to the input shaft 10. The crank pin 30 is
positioned at the rotational axis of the input disk 20, however, so
the rotation of the input shaft is not transmitted to the
connecting rod 40 or ultimately to the driveshaft 110.
[0035] FIG. 20 shows the continuously variable transmission as
illustrated in FIG. 19, but with the crank pin displaced slightly
from the rotational axis of the input disk. The rotational movement
of the input shaft is transmitted to the output shaft 110 at a
ratio of 10:1.
[0036] FIG. 21 shows the continuously variable transmission of
FIGS. 19 and 20 in which the crank pin is displaced further from
the axis of rotation of the input disk. The rotational movement of
the input shaft is transmitted to the driveshaft at a ratio of
1:1.
[0037] FIG. 22 shows an example of a continuously variable
transmission that is acting as an overdrive. The crank pin 30 is
positioned at the rotational axis of the input disk 20.
[0038] FIG. 23 shows the continuously variable transmission of FIG.
22 in which the crank pin 30 is displaced slightly from the
rotational axis of the input disk 20.
[0039] FIG. 24 shows the continuously variable transmission, again
acting in overdrive mode, in which the crank pin is displaced on
the face of the input disk further from the rotational axis of the
input disk than in the transmission shown in FIG. 23. The input
applied to the drive shaft 110 is 100 rpm counterclockwise, which
causes the output 100 to rotate counterclockwise at 1000 rpm.
[0040] FIG. 25 shows a continuously variable transmission having
reverse gear functionality.
[0041] FIG. 26 shows an embodiment of a continuously variable
transmission in which wheels drive an engine/flywheel at a 1:1
ratio. The crank pin is positioned on the face of the input disk 20
at the rotational axis of the input disk. A rotational input of
1000 rpm counterclockwise is applied through the driveshaft 110,
resulting in output to the engine/flywheel of 1000 rpm clockwise
(1:1 ratio).
[0042] FIG. 27 shows an embodiment of the continuously variable
transmission in which wheels drive an engine/flywheel at a 3:5
ratio. The crank pin is displaced 10 nm from the rotational axis of
the input disk.
[0043] FIG. 28 shows an embodiment of the continuously variable
transmission in which wheels drive an engine/flywheel at a 1:10
ratio. The crank pin is now displaced 30 mm from the rotational
axis of the input disk.
[0044] FIG. 29 shows an embodiment of the continuously variable
transmission in which an engine/flywheel drives wheels at a 0:1
ratio. The crank pin is displaced 33.3 mm from the rotational axis
of the input disk in this embodiment.
[0045] FIG. 30 shows an embodiment of the continuously variable
transmission in which an engine/flywheel drives wheels at a 10:1
ratio. A rotational force applied to the main backing plate 100,
with the crank pin positioned at 33.3 mm from the rotational axis
of the input disk (1/3 the bell crank radius). The 3:1 gearing
between the output shaft 70 and the drive shaft 110 results in a
zero rotational force applied to the driveshaft.
[0046] FIG. 31 shows an embodiment of the continuously variable
transmission in which a rotational input applied to the main
backing plate 100, when the crank pin is positioned 50 mm from the
rotational axis of the input disk (1/2 the bell crank radius),
results in the driveshaft rotates at 500 rpm in the direction
opposite to that in which the main backing plate rotates.
[0047] FIG. 32 shows an embodiment in which an engine/flywheel
drives wheels in reverse at a 1:1 ratio. The crank pin is
positioned 66.6 mm (2/3 the bell crank radius) from the rotational
axis of the input disk. In this configuration, a 1000 rpm
counterclockwise rotational force applied to the main backing plate
100 results in the driveshaft 110 rotating clockwise at 1000
rpm.
[0048] FIG. 33 shows a continuously variable transmission that
includes a forward/reverse shifting gear unit. This shifting unit
allows a single transmission to perform both overdrive and
underdrive duties, with the reversing output as observed in FIGS.
30-32 negated by the reversing mechanism.
[0049] FIG. 34 shows another embodiment of a transmission that
employs a forward/reverse shifting gear unit that allows the
transmission to perform both overdrive and underdrive duties.
[0050] FIG. 35 shows a wide ratio underdrive transmission that can
carry torque in both directions. In this configuration, the wheels
drive the engine/flywheel at a 1:1 ratio with the crank pin radius
at 0 mm.
[0051] FIG. 36 shows a wide-ratio underdrive transmission in a
configuration in which the crank pin radius is set at 66.6 mm (2/3
the bell crank radius) and the output shaft 70 rotates in the
direction opposite to that of the input. This reversal of the
rotation is itself reversed using a reversing gearbox that has a
ratio of 1:1. In contrast to the configuration shown in FIG. 35, in
this embodiment the engine/flywheel drives the wheels.
[0052] FIG. 37 shows a transmission in which the wheels drive the
engine/flywheel in overdrive mode. The crank pin radius is 16.66 mm
(less than 1/3 the bell crank radius). The wheels impart a
counterclockwise input of 1000 rpm to the driveshaft. This is
transmitted to the engine/flywheel at a 1:2 ratio (output to
engine/flywheel is 2000 rpm counterclockwise).
[0053] FIG. 38 shows a transmission in which the engine/flywheel
drives the wheels. A 2000 rpm counterclockwise input is applied to
the main backing plate when the crank pin is positioned at 50 mm
(1/2 the bell crank radius) from the rotational axis of the input
disk. The output shaft rotates at 1000 rpm clockwise (opposite
direction to the input applied to the main backing plate, at a 2:1
ratio). A reversing gearbox is used to reverse the reversed
intermediate output.
[0054] FIG. 39 shows a transmission in which the wheels drive the
engine/flywheel at a 1:10 ratio. The crank pin radius is 30 mm
(slightly less than 1/3 the bell crank radius of 100 mm). A
counterclockwise rotation applied to the main backing plate by the
engine/flywheel results in a clockwise rotation of the driveshaft
of 1000 rpm.
[0055] FIG. 40 shows a transmission configured so that the
engine/flywheel drive the wheels at a 1:10 ratio. The crank pin
radius is 36.66 mm (between 1/3 and 1/2 of the bell crank radius).
A 10,000 rpm counterclockwise input applied to the main backing
plate is converted to an intermediate output of 1000 rpm clockwise
(opposite direction) at the output shaft. The reversing gearbox
converts this back to a counterclockwise rotation.
[0056] FIG. 41 shows a transmission configured so that the wheels
attempt to drive the engine or flywheel at a 0:1 ratio. The crank
pin radius is one third the bell crank radius, and since the chain
ratio is 3:1, a counterclockwise input at the driveshaft 110 is
locked, while a clockwise input results in the transmission
freewheeling (no rotational input imparted to the engine or
flywheel).
[0057] FIG. 42 shows a transmission configured so that an input
applied to the main backing plate 100 (for example, by an engine or
flywheel) drives the final drive shaft 142 (which can be connected
to the wheels of a car) at a 1:0 ratio. The crankpin radius is set
at one third the bell crank radius and the two sprockets and chain
provide a 3:1 step-up in output to the driveshaft 110. For example,
a 10,000 rpm counterclockwise input applied to the main backing
plate results in a zero rpm clockwise output at the driveshaft. A
reversing gearbox 143 provides a final output of zero rpm
counterclockwise; the final drive shaft freewheels in the clockwise
direction.
[0058] FIG. 43 shows two transmissions that are mechanically
blended together into a single machine where they share a common
input, reaction (crank pin moving mechanism) and output.
[0059] FIG. 44 shows a transmission having a gear type drive system
with a typical overdrive to achieve a net 1:1 top ratio.
[0060] FIG. 45 also shows transmission having two output elements
that are connected using a gear type drive system; when operated in
a typical overdrive, this transmission achieves a net 1:1 top
ratio.
[0061] FIG. 46 shows an eight-element embodiment of a transmission
that uses a gear drive system to connect the output elements.
[0062] FIG. 47 shows an example of a continuously variable
transmission that has a chain and sprocket arrangement to connect
the two output elements.
[0063] FIG. 48 shows another example of a continuously variable
transmission that uses a chain and sprocket arrangement to connect
the four output elements.
[0064] FIG. 49 shows another example of a continuously variable
transmission that uses a chain and sprocket arrangement to connect
the eight output elements.
[0065] FIG. 50 shows example of a continuously variable
transmission that uses a simplified output gear system that can be
used in applications for which a reverse direction and a net
underdrive output are acceptable.
[0066] FIG. 51 shows a four-element version of a continuously
variable transmission that is suitable in applications for which
reverse rotation is not acceptable.
[0067] FIG. 52 shows an eight-element version of a continuously
variable transmission that is suitable in applications for which
reverse rotation is not acceptable.
[0068] FIGS. 53A-C show three different views of one type of
mechanism that is suitable for adjusting the position of the crank
pin 30 on the face of the input disk 20. The crank pin is shown
positioned at the rotational axis of the input disk 20, and the
adjustment mechanism can move the crank pin across the face of the
input disk.
[0069] FIGS. 54A-C show three different views of another example of
a mechanism that is suitable for adjusting the position of the
crank pin 30 on the face of the input disk 20.
[0070] FIG. 55 shows a mechanism that employs an electric motor to
adjust the crank pin 30 position on the face of the input disk
20.
[0071] FIG. 56 shows a mechanism that uses a hydraulic cylinder 170
and piston 160 to adjust the position of the crank pin 30 on the
face of the input disk 20.
[0072] FIG. 57 shows an example of a purely mechanical motion
translation device for adjusting the position of the crank pin 30
on the face of the input disk 20.
[0073] FIG. 58 shows two views of an example of a continuously
variable transmission in which multiple connecting rods 40 are
attached to a single crank pin 30.
[0074] FIG. 59 shows two views of another example of a continuously
variable transmission in which multiple connecting rods 40 are
attached to a single crank pin 30.
[0075] FIG. 60 shows an example of an over-run/reversing
mechanism.
[0076] FIG. 61 shows a variation on the overrun/reversing
mechanism.
[0077] FIG. 62 shows one embodiment of a complete continuously
variable transmission subunit.
[0078] FIG. 63 shows a basic automotive-type transmission in which
a CVT subunit, configured in underdrive mode, is coupled to an
over-run/reversing unit.
[0079] FIG. 64 shows another embodiment of an automotive-type
transmission in which a CVT subunit is coupled to an
over-run/reversing unit.
[0080] FIG. 65 shows a full automotive transmission system that
includes flywheel energy storage for regenerative braking.
[0081] FIG. 66 shows another example of a full automotive
transmission system that includes flywheel storage for regenerative
braking.
[0082] FIG. 67 shows the results of an efficiency test for a CVT
having eight output elements as described in the Example.
DETAILED DESCRIPTION
[0083] This invention provides a continuously variable transmission
(CVT). The transmission has pseudo-infinite ratio coverage. Speed
reduction can theoretically be varied from 1:1 to 1:0. Moreover, by
driving the output as an input and swapping the input and reaction
members, one can invert the transmission with resulting speed
ranges of 1:1 to 0:1. In this configuration the invention is an
overdrive CVT that theoretically can achieve an infinite overdrive
ratio. Very high overdrive ratios are practical with the
transmission.
[0084] The transmission described herein is a positive drive design
as opposed to a friction drive (belt, disk, ball) designs that were
previously known. The continuously variable transmissions of the
invention also lack movable fulcrums and levers that are found in
previously known transmissions.
[0085] In some embodiments, the continuously variable transmission
of the invention generally can carry torque in one direction at a
given ratio setting. Should the output want to run faster than the
input is driving it, the output will freewheel. For many
applications this is not a disadvantage. In automotive
applications, where engine braking is sometimes required, a fixed
ratio gear set can be clutched in at a desired ratio to provide
engine braking. Most current automatic transmissions also overrun
in their gear ranges and need additional hardware to provide engine
braking. This mechanism can be part of a reversing gear set that
all current CVT designs currently need and incorporate in their
practical designs. Also, a version of the transmission of the
invention is described herein that can carry positive torque and
negative torque. The switch from positive to negative torque can
require a change in set point of the crank pin radius that may take
a few seconds. Additionally, this version it permits the CVT to
operate as an underdrive transmission and an overdrive
transmission. Also, two separate transmission systems can be
blended into a single mechanical assembly to carry torque in a
bi-directional manner without time-consuming changes in operating
mode. These last two configurations are ideally suited as a
flywheel kinetic energy storage system for various vehicle
configurations that will be discussed as practical
applications.
[0086] Finally, the invention provides automotive transmission
systems with flywheel energy storage for vehicle braking as a
practical application of the continuously variable transmission.
Two such systems are described in detail. The first system
described requires separate underdrive and overdrive versions of
the transmission, but fulfills all automotive requirements
including reverse and overrun engine braking functions without
additional, specialized hardware. The second system only requires
one CVT unit configured for switching between both overdrive and
underdrive modes, but also requires a reversing gear unit.
Additionally, the switch between overdrive and underdrive modes in
the second system takes a finite amount of time (1-2 seconds)
versus immediate reaction in the system with separate CVT
units.
Continuously Variable Transmission--Principles of Operation
[0087] The transmission operates on the following principles of
operation. First, rotary motion is converted to variable amplitude,
linear reciprocating motion. The first element of this mechanism is
a rotating input shaft 10 that is operably connected to an input
disk 20 such that the input disk rotates about its center when the
input crankshaft rotates. A crank pin 30 is movably attached to a
face of the input disk such that the crank pin can be positioned
anywhere from the exact center of the input disk (FIG. 1A) to a
maximum position offset from center (FIG. 1B).
[0088] A connecting rod 40 is attached to the adjustable crank pin
30 as shown in FIG. 2, thereby converting the rotating motion of
the crank pin to a reciprocating linear motion. The amplitude of
this linear motion can vary from zero, when the crank pin is
positioned at the center of the input disk 20 (FIG. 2A), to twice
the maximum distance between the center of the input disk and the
position of the crank pin (FIG. 2B). The resulting continuously
variable transmission can continuously and infinitely vary its
output speed from zero to a maximum value. In doing so, it converts
a smoothly turning rotary motion to sinusoidal reciprocating
motion.
[0089] To convert this reciprocating motion to a smoothly turning
rotary motion whose speed is proportional to the amplitude of the
reciprocating motion, additional components that comprise an output
element can be attached to the above-described mechanism. The first
element of this output element is to attach the reciprocating end
of the connecting rod 40 to a bell crank 50 (FIG. 3). The rotating
axis of the bell crank is connected to a one-way clutch 60 (for
example, a sprag or roller clutch) that only transmits, for
example, the clockwise rotation of the bell crank to the output
shaft upon which it is mounted. When the bell crank oscillates in
the counterclockwise direction (for example), the one-way clutch
slips and does not transmit the counterclockwise rotation to the
output shaft. In FIG. 3A, the crank pin is located in the center of
the input disk, so rotation of the input disk does not result in
any movement of the connecting rod, the bell crank, or the output.
In FIG. 3B, the crank pin is displaced from the center of the input
disk. Rotation of the input disk in a clockwise direction (FIG.
3B-E) causes reciprocating motion of the connecting rod, which
causes the bell crank to oscillate. In the illustrated example, the
one-way clutch 60 transmits clockwise rotation to the output shaft
70, but does not transmit counter-clockwise motion due to slippage
of the one-way clutch.
[0090] It is important to note that the fixed length bell crank
should always be at least 20% longer than the maximum adjustment
distance of the adjustable crank pin from the center of the input
disk. If the distance of the crank pin from the center of the input
disk were to approach the length of the output bell crank, the bell
crank might cease the desired rotational oscillation and try to
circle around one way or the other and jam the machine. If the
crank pin were located a distance from the center of the input disk
that is greater than the length of the bell crank, the input would
jam on the first rotation. It is therefore necessary and desirable
to limit the maximum distance of the crank pin from the center of
the input disk to about 70% or 80% of the length of the bell crank
to ensure proper operation.
[0091] The device has now converted the reciprocating motion of the
crankshaft to a rotary motion whose average rotational speed is
proportional to the amplitude of the reciprocating motion. However,
the output rotational speed is not smooth. Not only does the output
speed vary over time, but also approximately 50% of the time (when
the bell crank is reciprocating in the freewheel direction) the
output shaft is not rotating at all. FIG. 4 shows an approximate
output rotational speed versus time diagram.
[0092] In order to create a smoother rotational output, additional
features can be added. The collection of parts consisting of the
connecting rod 40, bell crank 50 and the one-way clutch 60 is
defined herein as an "output element." If one adds a second output
element to the transmission, attached to the same crank pin 30 and
rotated 180 degrees around the axis of the crankshaft opposite the
first output element (FIG. 5). This second output element is
identical in function to the first except that it operates 180
degrees out of phase with the first. If the outputs of both
elements are then geared together in any convenient manner (chain
and sprockets, gear train, toothed belt and sprockets, belt and
pulley, etc.) each output element will then contribute output
rotation during the stationary period of the other. The result is a
much-improved output that while still having significant velocity
variations eliminates the large periods of stationary output. The
resulting output is graphically represented in FIG. 6. For a
limited number of applications this design may be sufficient and
acceptable.
[0093] If more output smoothness is desirable, two more output
elements can be added at 90 degrees and 270 degrees from the first
two elements, as illustrated in FIG. 7. The graphical
representation of the speed output of the four element transmission
is shown in FIG. 8. Four output elements overlapping lead to a much
higher level of output speed smoothness. For typical geometries,
the amount of speed variation is less than 10%. Coupled with some
type of output torsional dampening, (spring, friction, hydraulic,
slipping clutch, etc.) a four element CVT of this design could be
commercially acceptable for many applications, including automotive
applications.
[0094] If a premium application requires a very high level of
smoothness, an 8 element CVT can be implemented (FIG. 9). A
rotational speed versus time plot for this design is shown in FIG.
10. With typical geometries, speed variations with this design can
be reduced to about 2%. This should be sufficiently smooth for
almost all applications. Outputs with more elements can be built to
achieve even greater smoothness, but improvement is diminishing.
Many elements can also increase the load capacity of the
transmission as multiple elements begin to share the load in
parallel.
[0095] It should be noted that speed variations in practice should
be less than these theoretical numbers. The speed peaks will result
in marginally higher loading than the valleys. Normal elastic
deflections in all of the structural elements will provide built in
torsional dampening. These deflections will be higher during the
peak speeds--effectively slowing down the output slightly and
smoothing the output more.
[0096] Also note that the progression of examples being 1, 2, 4, 8
elements is not a requirement. Any number of output elements
including odd or prime numbers may be employed. Also, the spacing
around the crankshaft need only be evenly spaced if a uniform
output speed variation is desired. Practical design considerations
may prove that deviations from even spacing may prove to be
superior for the purpose of breaking up harmonic resonances in the
transmissions structure.
Alternative Hook Ups and Geometries of the Basic System
[0097] FIG. 9 shows one embodiment of the basic system having eight
output elements. Other arrangements and geometries can be viable
designs, which achieve different results.
[0098] One area to economize on is the number of rotating axes. The
system illustrated in FIG. 9 has nine--a large amount of
complexity. FIGS. 11 and 12 show alternate versions of the two
element and four element embodiments that are illustrated in FIGS.
5 and 7, respectively. This embodiment reduces the number of axes
approximately in half. The elements that did reside 180 degrees out
of phase on separate axes now reside on the same shaft, each with
its own one-way clutch. If the connecting rods were infinitely
long, these designs would be geometrically identical to those
illustrated in FIGS. 5 and 7. But given the angularity of finite
length connecting rods, these two bell cranks do not act on the
output shaft exactly 180 degrees apart. The phase shift between
bell crank speed contributions would alternate between slightly
less than 180 degrees to slightly more than 180 degrees. This may
not be detrimental for all applications and might even impart an
advantage from a noise and vibration standpoint as discussed
earlier.
[0099] This concept of pairing elements on common shafts can be
extrapolated to an even greater number of elements than discussed
so far. Any number can be incorporated as long as the elements are
engineered to not interfere with one another in the final design
package.
Scotch Yoke Designs
[0100] Instead of the adjustable crank pin driving the bell crank
elements through a pivoting connecting rod, a whole family of
designs can be created whereby the crank pin 30 drives the bell
cranks via a slot 80 in a scotch yoke 90 (FIGS. 13 and 14).
Although scotch yokes typically suffer from higher Hertzian
stresses at the crank pin roller/slot interface this can be
alleviated through the use of linear roller bearings in the
slot.
[0101] The main advantage of these concepts is better geometries
with less translating errors. This is illustrated best in the
three-axis design of FIGS. 15 to 18. In FIG. 15, one can see that
the scotch yoke maintains its parallelism throughout its range of
motion--leading to a minimal geometrical error (and least speed
variations) relative to the other proposed designs. FIGS. 16 to 18
illustrate another major advantage of the scotch yoke. Unlike a
connecting rod, it is not necessary to create a new axis rotated at
a fixed angle to create a new element with a different phase
output. All that is necessary is to rotate the slot in the scotch
yoke and adjust the phasing of the output bell cranks. For example,
in FIG. 15 the slot 80 is vertical, the slotted link reciprocates
left to right and the power stroke occurs when the crankpin is in
the vicinity of 12 o'clock and 6 o'clock. In the embodiment shown
in FIG. 16, the slot 80 is horizontal, the slotted link transmits
power by moving up and down, the bell cranks are rotated 90 degrees
and the power stoke occurs when the crankpin passes through 3
o'clock and 9 o'clock. In this arrangement, the bell cranks must be
rotationally synchronized by a secondary bell crank and link system
that includes a secondary link 93 that connects a first secondary
bell crank 94 and a second secondary bell crank 95. Without this
synchronizing link, the slotted link would pivot around the
crankpin during the power stroke and the opposite end would
teeter-totter and push the opposite bell crank in the freewheel
direction. The end result is that force could never be transmitted
in the desired direction. By incorporating the synchronizing link
system, the slotted link is forced to remain parallel during
operation and power is effectively transmitted to the appropriate
bell cranks. The same discussion pertains to the element systems
shown in FIGS. 17 and 18 except there the slot and power stoke
occur plus and minus 45 degrees from the vertical. The slotted
links translate in a linear motion plus and minus 45 degrees from
the vertical and the output bell cranks are indexed 45 degrees
appropriately. A synchronizing link system is also necessary in
both of these arrangements for the identical reasons stated for the
system shown in FIG. 16.
[0102] If one were to stack up FIGS. 15-18 into one machine on a
common crankshaft, an eight element CVT can be created with just
two output axes. These two outputs rotate in opposite directions so
appropriate measures must be taken when gearing them together. (See
later section on output gearing options).
Overdrive Version of CVT
[0103] The theory of operation of the CVT has been described herein
and is most easily understood as an underdrive transmission. That
is, for a fixed input speed, say 1000 rpm, ratio changing can vary
the output speed from 1000 rpm (1:1) to zero rpm (1:0). This is
illustrated in the three dimensional diagrams featured in FIGS.
19-21. In these diagrams, only one output element is shown for
simplicity.
[0104] FIG. 19 specifically shows the case with the crank pin 30
attached to the input disk 20 on the rotational axis of the
rotating input shaft 10, so the crank pin spins freely inside the
connecting rod 40 with no reciprocating action. The resulting
output is therefore zero (1:0). In the embodiments shown in FIGS.
19 to 21, the main backing plate 100--to which all the output
elements are mounted via bearings--serves as the main reaction
member. This plate is grounded to the transmission's case and is
always at zero speed. Optionally, the output shaft 70 can be
connected to a driveshaft 110 by means of belts, chains and
sprockets, gears, or other means known to those of skill in the
art. In the illustrated embodiment, a first sprocket 120 drives a
second sprocket 130 by means of a belt or chain 140. Gear trains or
other mechanisms are also suitable for use in transmitting the
rotation of the output shaft to the driveshaft 110.
[0105] FIG. 20 shows the transmission in which the crank pin
adjusted slightly off of the center of the input disk 20, resulting
in a 10 to 1 speed reduction (underdrive) (10:1). The connecting
rod 40 moves back and forth, causing reciprocating clockwise and
counterclockwise rotational movement of the bell crank 50, with the
clockwise movement (in the illustrated example) being translated by
the one-way clutch 60 into rotation of an output shaft 70. As in
FIG. 19, a pair of sprockets and a chain or belt are illustrated
driving a driveshaft 110.
[0106] FIG. 21 is the 1:1 ratio case. Here the crank pin 30 is at
its maximum design radius from the rotational axis of the input
disk 20 (approximately 70% of the bell crank length). The sprockets
120 and 130 and associated chain or belt inverts the same ratio
(1/0.70=143%) so that the resulting driveshaft 110 speed is the
same as the input speed.
[0107] By switching the input, output and reaction elements, an
overdrive version of the transmission can be created. This is best
understood by comparing FIGS. 20 and 22. The transmissions are in
the same configuration (crank pin 30 on axis of input disk 20) but
they differ by reference frame. FIG. 22 shows a reference frame 100
rotating around the transmission at 1000 rpm. The net effect is to
subtract 1000 rpm from the input shaft, the main support plate and
the output shaft. What was the input shaft 10 in FIG. 20 is now at
zero speed and becomes the new reaction member. What was the output
(driveshaft) 110 in FIG. 20 is the new input member and is now
rotating counterclockwise at 1000 rpm. What was the reaction plate
100 is now the output and is rotating counter clockwise at 1000
rpm. So in the new overdrive configuration, this is the 1:1 ratio
case.
[0108] In FIG. 23, the crank pin 30 is adjusted slightly off of
center of the input disk 20. This results in a slight overdrive.
With the new input 110 rotating at 900 rpm, the new output 100 is
turning at a 1000 rpm. If the input 110 were accelerated to 1000
rpm, the output would spin at 1000/900.times.1000=1111 rpm.
[0109] In FIG. 24, the crank pin 30 is adjusted further out from
the rotational axis of the input disk and the transmission now has
a 10.times. overdrive ratio. With an input 110 speed of 100 rpm,
the output 100 spins at 1000 rpm. Likewise, if we were to spin the
input 110 at 1000 rpm, the output 100 would spin at 10,000 rpm. As
the crank pin 30 is adjusted further out, the output will
theoretically attempt to achieve infinite speed. This cannot be
achieved in practice, as friction will cause the input member to
lock up at some finite ratio. However, quite large finite overdrive
ratios should be practical. This configuration would result in a
device that is uniquely capable of accelerating a flywheel by
purely mechanical means for the purpose of energy storage.
Overdrive CVT with Reversing Capability
[0110] With a slight modification, the overdrive configuration can
be employed to give useful reverse gear functionality for certain
transmission configurations.
[0111] FIG. 25 illustrates the overdrive CVT configuration but with
a greater step up in the output gearing than what is necessary to
compensate for the maximum lever ratios of the crank pin/bell
crank. In the illustrated example, this output ratio is a 1:2 step
up in speed. When the distance of the crank pin 30 from the
rotational axis of the input disk 20 (this distance is termed the
"crank pin radius" herein) reaches one half the radius of the fixed
bell crank (the distance between the rotational axis of the output
shaft 70 and the center of the attachment point of the connecting
rod to the bell crank; this distance is termed the "bell crank
radius" herein), the transmission has reached the theoretical
infinite speed overdrive. The transmission should only pass through
this point with the output not turning or declutched from the
driveline.
[0112] If the crank pin is adjusted to a radius greater than 50% of
the bell crank radius--say to the maximum design radius of 70%- and
the input and outputs are reversed, the transmission will reverse
output rotation. Specifically, if one drives the support plate 100
in the same direction (as shown), the geared output will rotate in
the opposite direction with a speed reduction. In this case with
1000 rpm on the support plate 100, the result would be a 400 rpm
reverse rotation on the output gearing.
[0113] Exploiting this ability to reverse in a practical automotive
transmission design is discussed below.
Combination Overdrive/Underdrive CVT
[0114] If the step up in the output gearing is changed to 1:3, the
transmission still operates as an overdrive transmission as before.
However this time it reaches the theoretical infinite overdrive
point when the crank pin radius reaches one third of fixed bell
crank radius. Moving beyond that radius, the output reverses, as in
the prior configuration. And like before, the output rotates at an
underdrive ratio with the main support plate driven as an input
relative to the output. Right at the one-third crank pin radius
point the underdrive is infinite from the main support plate to the
output (1:0). That is, the main support plate 100 as the input can
be freely rotated with no transfer of rotation to the output. As
the crank pin radius is made slightly greater than one third of the
fixed bell crank radius, the output will start to turn slowly
relative to the main support plate input with a maximum underdrive
reduction. As the crank pin radius increases out to two thirds of
the bell crank radius, the underdrive ratio increases up to
1:1.
[0115] FIGS. 26-32 illustrate various speed settings of both
overdrive and underdrive modes. In the embodiment shown in FIG. 26,
wheels drive the driveshaft 110 counter-clockwise at 1000 rpm. A
pair of sprockets 120 and 130 and a chain 140 drive the output
shaft 70 at a 1:3 ratio. With the crank pin radius set at zero mm
from the rotational axis of the input disk and the bell crank
radius at 100 mm, the output to the engine/flywheel 100 is the same
as the input (1000 rpm counterclockwise).
[0116] In FIG. 27, the input is reduced to 600 rpm
counterclockwise. But by setting the crank pin radius at 10 mm, the
output 100 remains at 1000 rpm counterclockwise. In FIG. 28, the
input is reduced still further, to 100 rpm counterclockwise. By
moving the crank pin to 30 mm from the rotational center of the
input disk, the output is maintained at 1000 rpm
counterclockwise.
[0117] FIG. 29 shows the transmission with the crank pin radius set
at one third the bell crank radius. Now, the input force (1000 rpm
counterclockwise) is applied from the engine/flywheel to the main
backing plate 100. The chain ratio is set at 3:1, thereby reducing
the output to the driveshaft 110 to zero.
[0118] When the crank pin radius is increased beyond one-third of
the bell crank distance as shown in FIG. 30, the transmission
converts the 1000 rpm counterclockwise input to a 100 rpm clockwise
(the opposite direction of the input) output at the driveshaft 110.
A further increase in the crank pin radius, as illustrated in FIG.
31, results in a 1000 rpm counterclockwise input being converted to
a 500 rpm clockwise output at the driveshaft 110. Increasing the
crank pin radius to two-thirds the bell crank radius as shown in
FIG. 32 results in the 1000 rpm counterclockwise input being
reversed to a 1000 rpm clockwise output at the driveshaft 110.
[0119] A CVT configured as described above can therefore cover a
theoretically infinite ratio range from underdrive to
overdrive--the only drawback being the reversal of output direction
as the CVT transitions from overdrive to underdrive modes. If a
forward/reverse shifting gear unit is connected to the output of
this transmission (FIGS. 33 and 34) and it is reversed as the
transmission transitions from overdrive to underdrive mode, the
transmission can perform both overdrive and underdrive duties with
the reversing output being negated by the add on reversing
mechanism. With the output of the reversing gear unit connected to
the drive wheels of a vehicle, and the main support plate connected
rotate-ably to an energy storage flywheel, this configuration can
serve both overdrive and underdrive functions for a vehicular
flywheel energy storage system. This system is discussed in greater
detail below.
Underdrive CVT with Bi-Directional Torque Carrying Capability
[0120] Two versions of the above transmissions can be created and
designed into a single transmission machine to create a wide ratio
underdrive transmission that can carry torque in both directions.
FIG. 35 shows a transmission with the crank pin radius (distance of
crank pin 30 from rotational center of input disk 20) set at zero
mm, the bell crank radius is 100 mm and the final drive ratio set
at 3:1. An input force of 1000 rpm counterclockwise is applied to
the driveshaft 110 (e.g., by the wheels of a car). The chain and
sprockets reduce this by 3:1 in this example, and the output to the
engine/flywheel is also 1000 rpm counterclockwise. In FIG. 36), the
crank pin radius of the transmission is set at 66.6 mm (two-thirds
the bell crank radius, which is again 100 mm). Now, the input force
is applied to the main backing plate 100 (for example, by the
engine or flywheel of a car). In the illustration, the input force
applied by the engine or flywheel is 1000 rpm counterclockwise,
resulting in a clockwise rotation of the output shaft 70. The chain
and sprockets step up the rotation of the driveshaft 110 to 1000
rpm, but still a clockwise direction. To compensate for the
reversal in direction, a reversing gearbox 143 is employed. In the
illustrated example, the reversing gearbox has a 1:1 drive ratio,
resulting in the output at the final drive shaft 142 being 1000 rpm
counterclockwise. Thus, the transmission can transmit torque in
either direction--from the engine/flywheel to the wheels, or from
the wheels to the engine/flywheel.
[0121] FIGS. 37 and 38 show a continuously variable transmission
with its crank pin radii at 16.6 mm and 50 mm, respectively. The
bell crank radius is 100 mm, and the chain ratio is 3:1. In the
example shown in FIG. 37, torque is transmitted from the wheels
(input at the drive shaft 110 is 1000 rpm counterclockwise) through
the sprockets and chain to the output shaft 70. With the crank pin
radius set at less than one third the bell crank radius, the
one-way clutch transmits a counter-clockwise rotation of 2000 rpm
to the main backing plate 100. In the example shown in FIG. 38, the
input force is applied to the main backing plate 100 (e.g., by an
engine or flywheel of a car) instead of to the driveshaft as in
FIG. 37. A 2000 rpm counterclockwise input results in a clockwise
rotation of the output shaft 70. The 3:1 ratio provided by the
chain and sprockets step the output up to 1000 rpm clockwise at the
driveshaft 110. To convert the output to a counterclockwise
rotation at the final drive shaft 142, a reversing gearbox 143 is
employed.
[0122] FIGS. 39 and 40 again show the continuously variable
transmission, this time with their crank pin radii at 30 mm and
36.6 mm, respectively, and a bell crank radius of 100 mm in each
case. The transmissions provide a 10:1 reduction with torque
transmitted in both directions (in FIG. 39, torque is transmitted
from the driveshaft 110 to the main backing plate 100 (e.g., to the
engine/flywheel), while in FIG. 40, torque is transmitted from the
main backing plate to the driveshaft). In FIG. 39, a 1,000 rpm
counterclockwise input results at the driveshaft results in a
10,000 rpm counterclockwise output at the main backing plate 100.
In FIG. 40, a 10,000 rpm counterclockwise input at the main backing
plate results in a final output of 1000 rpm counterclockwise
(through the reversing gearbox 143 that is attached to the
transmission).
[0123] When both crank pins converge at 33.3 mm (one third the bell
crank radius) (FIGS. 41 and 42), both transmissions have achieved
0:1 speed ratio. The engine/flywheel is free to spin at any speed
without transmitting any torque and the wheels are locked, unable
to transmit torque in either direction.
[0124] These two transmissions (one with a reversing gearbox and
one without) can be mechanically blended together into a single
machine where they share a common input, reaction (crank pin moving
mechanism) and output. An example of one such embodiment is shown
in FIG. 43. Two independent sets of bell cranks and output gear
systems are combined into a single rotatable assembly (input). The
two systems are attached to separate crank pins, which are in turn
attached to a common sliding mechanism (reaction). The two crank
pins are permanently spaced 66.6 mm apart. The sliding mechanism's
range of motion is from a point where one crank pin is at the
center of the machine with the other is 66.6 mm off center, to a
point where both crank pins are 33.3 mm offset on opposite sides of
the center. The two output gear systems transmit motion through
separate, coaxial shafts. (One solid shaft, one hollow shaft) The
hollow shaft enters the input of the reversing gearset and the
solid shaft passes through the reversing gearset and is permanently
(without a clutching mechanism) and rotatably connected to the
output of the reversing gearset (Output).
Output Gearing Options
[0125] In typical versions of CVT of the invention, the various
numbers of output elements are geared positively together in a
rotational sense. This can be accomplished with chains and
sprockets, gears, toothed belts, etc., or any other positive type
of drive mechanism. FIGS. 44 and 45 show a gear type drive system
with a typical overdrive CVT to achieve a net 1:1 top ratio. Two
output element (FIG. 44) and four output element (FIG. 45)
embodiments are shown. For simplicity, the pitch diameters of
meshing gears is represented by the circles drawn concentric with
each axis. Where the circles touch tangentially is where two gears
mesh with each other. FIG. 46 illustrates an eight-element version.
In each of these arrangements, all of the bell crank/freewheeler
elements can transmit rotational power to the center output shaft.
The overdrive output ratio with eight elements as shown in FIG. 46
has drive gears so large and close together that they overlap each
other. This can be mechanized by incorporating two rows of gearing
so that the drive gears alternate axially with each other. These
two staggered rows of four drive gears then mesh with two stacked
driven gears on the common output shaft 70. All of these
embodiments reverse the output direction from the input.
[0126] FIGS. 47 to 49 show chain and sprocket arrangements. Again,
for simplicity lines tangentially intersecting circles represent
the chains. The circles represent the pitch diameters of the
sprockets. Chains require a certain amount of sprocket wrap angle
so all versions require multiple, staggered output sprockets.
Anywhere chains have to cross one another would require a new row
of sprockets axially offset along the respective shaft to clear the
other chain. Two and four axis arrangements (FIGS. 47, 48) require
two staggered rows of sprockets while the eight axis arrangement
(FIG. 48) requires four staggered rows of sprockets. Chains do not
reverse direction.
[0127] If a reverse direction and a net underdrive output are
acceptable, a greatly simplified output gear system can be employed
as shown in FIG. 50. Using a large driven gear and small driven
gears permit the whole gear train to be package on a single plane
of gears.
[0128] If reverse rotation is not acceptable and a premium gear
drive with a large output overdrive (1:3) is needed, then an
intermediate idler gear can be employed to mesh two outer gears and
reach a small final driven gear. FIGS. 51 and 52 show four- and
eight-output element embodiments respectively.
Design Examples for Certain Aspects of the CVT
[0129] Moveable Crank Pin on Input Crankshaft
[0130] In some embodiments, for example as shown in FIG. 1, the
crank pin 30 is mounted onto a flat plate 145 that engages slots in
the face of the input disk 20. The crank pin is mounted on one end
of the plate 145 so that that at one end of its sliding range, the
crank pin lies in line with the rotational axis of the input
crankshaft (FIG. 1A). At the other extreme of its travel, the crank
pin is at its maximum design radius (FIG. 1B). The crank pin/plate
assembly could also include a threaded hole along its sliding axis
and could be moved through its range of motion by, for example,
rotating a threaded rod, mounted inside the disk with a bearing at
one or both ends.
[0131] Alternatively, the same crank pin sliding plate assembly 145
could be designed with a protruding dog element 155 in its
underside instead of a threaded hole (FIG. 53). This protruding dog
would engage a spiral wheel that can be rotated on the same axis as
the input shaft. As the spiral wheel rotates the dog on the crank
pin assembly would move in a radial direction, moving the sliding
crank pin assembly through its full range of motion.
[0132] Alternatively, the crank pin 30 could be mounted in an
offset from center location on a rotatably moveable disk 150 (FIG.
54). This disk would be smaller than and embedded in an offset
position on the input disk 20. As this crank pin disk 150 is
adjustably rotated through 180 degrees, the crank pin can be
positioned from the center of the input crankshaft 10 to the
designed maximum radius or anywhere in between. An external gear on
the back of the crank pin disk can engage an adjusting spur gear
for positioning. (See adjustment mechanisms discussion below).
[0133] Crank pin adjustment in the stationary reference frame
[0134] In some of the embodiments of the movable crank pin
mechanisms, the mechanisms reside on the input crankshaft, which is
typically rotating at high speed. The crank pin position needs to
be adjusted in a controlled manner to a defined position
independent of the input system's rotating conditions.
[0135] In one embodiment, an electric motor of any type (servo,
stepper, DC, AC induction, etc.) is mounted on the input crankshaft
10 and is rotatably connected to the adjustment mechanism by means
well known to those skilled in the art (e.g., gears, chain and
sprockets, couplers, etc.) The electrical connections to this motor
would be transferred to the stationary reference frame via
electrical slip rings and brushes (FIG. 55).
[0136] In another embodiment, the sliding crank pin 145 adjustment
is positioned via a hydraulic actuator consisting of a piston 160
and cylinder 170 (FIG. 56). This double acting actuator is moved by
high-pressure hydraulic fluid pressure that communicates with the
stationary reference frame via rotating hydraulic seal rings. These
seals are well known art in hydraulic system design.
[0137] There may be applications where the prime mover assigned to
adjust the ratio of the CVT cannot be limited to devices with power
and control connections that are readily adaptable to translation
through a rotating reference frame. For these applications, purely
mechanical motion translation devices are provided. One such device
is shown in FIG. 57. Here, the adjustment mechanism is geared via a
set of spur gears, radial shaft and spiral bevel gears to an
adjustment sleeve 180 supported by a needle bearing coaxial with
the input shaft 10. When this sleeve is rotated relative to the
input shaft, the crank pin adjustment mechanism 145 responds.
[0138] Translating the relative rotation of this sleeve from the
rotating to stationary reference frames is accomplished through the
novel connection of two identically ratioed planetary gear sets
(FIG. 57, indicated by A and B). The connections of these planetary
gear sets is as follows:
[0139] Sun Gear A--Adjustment Sleeve 180
[0140] Ring Gear A--Case Ground 190
[0141] Sun Gear B--Input Shaft 10
[0142] Ring Gear B--External Adjustment Mechanism 200
[0143] Planet Carrier A--Planet Carrier B
[0144] It can be seen that if ring gear B is held stationary like
ring gear A, all elements of each gear set will turn at the same
speeds as the other gear set. Another way to say this is since sun
gear B is connected to the input shaft, both planet carriers are
tied together and both ring gears are stationary therefore sun gear
A and the adjustment sleeve will also turn precisely at the same
speed as sun gear B and the input shaft. Since there is no relative
motion between the input shaft and the adjustment sleeve--no change
in the crank pin offset will occur.
[0145] Now take the case where the input shaft is stationary. If
ring gear B is turned, planet carrier B will also turn somewhat
slower depending on the ratio of the sum of the ring and sun gears
number of teeth divided by the number of ring gear teeth (S+R)/R.
Planet carrier A will turn in concert with planet carrier B. Since
ring gear A is always grounded, sun gear A (and the adjustment
sleeve) will turn somewhat faster with a ratio to the planet
carriers depending on the number of teeth of the sun gear divided
by the sum of the number of teeth of the sun gear and ring gear
S/(S+R). The net result of this is the adjustment sleeve will turn
with a relative ratio to ring gear B depending on the number of
teeth of the sun gears divided by the number of teeth of the ring
gears (S+R)/R.times.S/(S+R)=S/R.
[0146] The crank pin offset can therefore be adjusted by turning
ring gear B relative to the stationary reference frame. This can be
accomplished independently of the rotational speed of the input
crankshaft.
[0147] Attachment of Multiple Connecting Rods to a Single Crank
Pin
[0148] A continuously variable transmission of the present
invention can have multiple connecting rods attached to a single
crank pin. A basic embodiment is shown in FIG. 58. Here, eight very
thin but broad connecting rods 40 are simply stacked up next to one
another on a single needle roller bearing 210. It should be noted
that it is generally desirable to keep the length of the crank pin
as short as possible to minimize the amount of cantilevered load
and consequent deflection on the crank pin. In the ideal situation,
a separate needle bearing should support each connecting rod. This
is due to the fact that the connecting rods undergo a slight
rotational displacement relative to one another during operation.
Implementing a separate needle bearing for each connecting rod
would force the crank pin to be excessively long. With a single
needle bearing supporting all of the connecting rods some slight
skidding of the rollers will occur, but since typically only one of
the connecting rods is heavily loaded at a time, any skidding will
occur on unloaded connecting rods and should not cause a problem
with excessive wear.
[0149] Another particular embodiment of the crank pin attachment is
shown in FIG. 59. Here a short intermediate hub 220 supported by a
single needle bearing 210 is mounted on the crank pin 30 and can
freely rotate. One connecting rod 230 is rigidly attached to the
hub. The remaining connecting rods 40 are connected to the
perimeter of the hub via freely swiveling pin connections 240. This
way, each connecting rod can freely move in their required slight
angular swiveling relative to one another, while the major
rotational motion is borne by the single needle bearing and hub
assembly.
[0150] Dedicated Over-Run/Reversing Mechanism
[0151] One embodiment of an over-run/reversing mechanism is shown
in FIG. 60. This mechanism is suitable for use in, for example, the
transmission shown as 143 in FIG. 36. The mechanism consists of a
single planetary gear set, a rotating clutch and one stationary
band clutch. The two clutches can be hydraulically operated as is
typical in automatic transmission design, although other mechanisms
for operating clutches as are known in the art are also suitable.
The input from the engine, in addition to driving the input of the
CVT, is connected in parallel to the sun gear of the planetary gear
set. The output of the CVT is connected in parallel to the ring
gear of the planetary gear set. The planet carrier of the gearset
is connected to both the rotating clutch and the band clutch.
During CVT operation, both clutches are released permitting both
the input and output to spin freely and independently. For direct
drive (engine braking mode) the rotating clutch would be engaged,
locking the planet carrier to the input shaft. This results in a
1:1 direct drive condition. For reverse operation, the band clutch
would be engaged with the rotating clutch disengaged. This would
ground the planet carrier causing the output to rotate in the
opposite direction of the input sun gear. Additionally, the output
speed would be reduced from the input speed proportional to the
number of gear teeth in the ring gear divided by the number of
teeth in the sun gear. R/S
[0152] Forward/Reversing Mechanism with 1:1 Ratio in Forward and
Reverse
[0153] FIG. 61 shows a variation on the overrun/reversing
mechanism. Here the planetary gearset is replaced by a spiral bevel
differential gearset. This mechanism has the same kinematics as the
planetary reverser except the differential gearset has no speed
change in the reverse mode. This type of reverser is desirable for
the Combination Overdrive/Underdrive CVT that is described in more
detail below.
Basic CVT Subunit
[0154] FIG. 62 is one embodiment of a complete CVT subunit. Here
the main backing plate 100 for the output elements is consolidated
into a rotatable unit that can transmit drive rotation externally
through a hollow shaft (labeled "reaction" in FIG. 62). The input,
reaction and output are labeled in FIG. 62 for the underdrive case,
but as discussed earlier these can be interchanged in order to
create an overdrive CVT. In underdrive mode, the application of a
rotational force to the input 10 (e.g., by the motor or flywheel of
a car) causes rotation of the input disk 20. Depending on the
relationship between the crank pin 30 radius, the bell crank
radius, and the ratio provided by the first sprocket 120 and second
sprocket 130, an output is provided to the driveshaft 110. If
instead a rotational force is applied to the output 110 (e.g., by
the wheels of a car), the transmission can act in overdrive mode to
impart a rotational force to the motor or flywheel as discussed
herein.
Basic Automotive Automatic Transmission
[0155] In FIG. 63, the basic CVT subunit, configured in underdrive
mode, is coupled to the over-run/reversing unit to create a simple
but practical automotive type CVT. The over-run/reversing unit 220
is connected mechanically in parallel to the CVT subunit. In most
forward driving conditions the reversing unit is in neutral. The
application of a rotational force to the input shaft 10 when the
crank pin 30 is displaced from the axis of rotation of the input
disk 20 imparts a rotational movement to the connecting rods 40,
which impart rotational movement to the output shafts 70 through
the bell crank and the one-way clutch as described above. Gears,
chains and sprockets, or a similar mechanism connect the output
shaft to the driveshaft 110, thereby imparting a rotational force
to the driveshaft. When engine braking is required, the rotating
clutch would be engaged, thereby transmitting rotational force from
the driveshaft 110 to the input 10. It should be noted that engine
braking, in this particular example, is limited to a single
ratio--in this case 1:1. If reverse gear is desired, the band
clutch would be engaged, the rotating clutch disengaged and the CVT
must be in the 0:1 ratio state. If the CVT were in a finite forward
ratio where the transmission attempts to reverse it would lock
up.
[0156] Another embodiment of a basic automotive transmission is
shown in FIG. 64. Here, the over-run/reversing unit 220 is the
spiral bevel differential gearset shown in FIG. 61. A worm gear
drive coupled to a rack and pinion drive is used to adjust the
position of the crank pin 30 on the face of the input disk 20.
Hybrid Kinetic Powertrain System
[0157] The invention provides a novel, vehicle based powertrain
system that can theoretically capture most of the kinetic energy of
a vehicle during braking and then use that energy with high
efficiency during subsequent vehicle accelerations. The energy
storage system can capture and use most or all of the kinetic
energy normally lost during braking, thereby allowing very large
efficiencies to be realized. Using this invention, a practical
vehicle can be constructed with an internal combustion engine
downsized all the way to 10 or 20 HP. To do this, the small engine
initially charges up the high efficiency storage system with enough
energy to propel the vehicle up to 100 km/hr. The vehicle then
accelerates to speed using this temporarily stored energy. Once at
cruising speed, the small gas engine maintains the vehicle's
velocity. When the vehicle needs to be braked to a stop, the
majority of the vehicle's kinetic energy is stored in the storage
system. When time to reaccelerate, that power comes from the
storage system. The small gas engine is only used to overcome
rolling resistance, aerodynamic resistance and any losses in the
turnaround efficiency of the storage system.
[0158] The invention described herein makes such a system practical
by providing a system that can capture and reuse the kinetic energy
of a vehicle currently lost during braking with no energy
conversions steps. The linear kinetic energy of the vehicle to is
transformed to rotational kinetic energy in a flywheel energy
storage system. To reuse this stored energy, the rotational kinetic
energy of the flywheel is transformed back into linear vehicle
kinetic energy, thereby re-accelerating the vehicle back up to near
its original speed.
[0159] The continuously variable transmission provided by the
present invention overcomes the two deficiencies of the hybrid
electric vehicle to exploit regenerative braking for energy
savings. First is overall turn-around efficiency. If the CVT is 90%
mechanically efficient, the overall turn around efficiency would be
(0.9).sup.2=81%. Secondly, there is no physical limit to the rate
at which a flywheel can accept energy flux. One need only
accelerate it harder.
[0160] A vehicle described above with only a 10 or 20 HP engine
would be operating at maximum load and efficiency all the time it
is operating. Plus, because all of the braking energy is reused for
accelerations, there would never be a situation where the engine is
burning fuel at a 70 or 100 HP rate during accelerations. The small
engine absolutely limits fuel consumption by its size alone.
[0161] The CVT that enables this system requires a very large ratio
coverage to handle the extremely wide speed swings this system must
accomplish to transform the kinetic energy in operation.
[0162] To accomplish energy transformation during braking, the
flywheel, which initially would be spinning at a relatively low
speed, would have to be accelerated by the output of the CVT to a
very high speed of about 10,000 to 20,000 rpm. All the while, the
input to the CVT from the vehicles driveline would be decelerating
from an initial high speed to a relatively low speed. This
necessitates a CVT with ratio coverage of 100:1 or more.
[0163] Likewise, to accomplish energy transformation during
re-acceleration, the output of the CVT to the vehicle's driveline
would have to accelerate from a relatively low speed to its final
high speed. While this is occurring, the flywheel, which initially
would be spinning at 10,000 to 20,000 rpm, would need to be
decelerated by the input of the CVT to a relatively low speed.
[0164] The transmission design proposed is capable of this extreme
ratio coverage. It has a theoretical range of infinity, but of
course this is not possible in practice. However, ratio ranges in
the hundreds are practical with mechanical efficiencies in the 90%
range.
[0165] The CVT inventions provided herein provide a practical
automotive transmission system that theoretically can capture most
of the kinetic energy of a vehicle during braking and then use that
energy with high efficiency during subsequent vehicle
accelerations. This is referred to as regenerative braking. Many
electric and hybrid electric vehicles make some attempt at
regenerative braking, but with very limited results. There are
several reasons for this. First the multiple energy conversions
steps that must occur in electric systems hamper the turn-around
efficiency. In an electric or hybrid electric vehicle the vehicles
kinetic (mechanical) energy must first be converted to electrical
energy in the motor/generator. Then this electrical energy must be
converted to chemical energy in the battery systems. The systems
power electronics also introduce losses. Then all of the losses in
each of these conversion steps is repeated and compounded when it
is time to convert the energy back to a usable form to accelerate
the vehicle. This amounts to a very large net loss in efficiency.
Then there is the problem of battery charging capacity. Even if the
conversion losses were not so large, the amount of energy released
during the rapid deceleration of a three thousand pound vehicle is
more than any battery system can hope to capture. Batteries just
can't be charged that fast. Some work is being done to develop
ultra capacitors to accept the energy rapidly, but with an energy
density approximately 400 times worse than batteries, an ultra
capacitor big enough to accept all the energy from just one
deceleration would be too large to install in a vehicle.
[0166] In the design described herein, vehicle braking energy is
captured via an overdrive CVT into a flywheel. There are no energy
conversion losses as the energy remains in the mechanical kinetic
form at all times. The only losses would be mechanical bearing
friction, which is typically quite low. There is also no rate limit
on energy capture as it simply amounts to accelerating the flywheel
at a faster rate. Also, since the flywheel need only store enough
energy for one or two complete accelerations it can be of modest
size and weight and be made of high strength steel rather than more
exotic materials.-If a mechanical flywheel regeneration system can
efficiently capture and return most of the energy from stopping a
vehicle back into accelerating it back up to speed, the vehicles
fuel-consuming engine could be downsized drastically. Theoretically
it would be possible to downsize the engine to a size only capable
of maintaining cruise velocity. This would translate to an 80 or 90
percent reduction in size, with tremendous gain in fuel
economy.
Full Automotive Transmission System with Flywheel Energy Storage
for Regenerative Braking
[0167] The invention also provides automotive transmission systems
that have flywheel energy storage for regenerative braking.
[0168] Underdrive/Overdrive CVT Combination Automotive Transmission
System
[0169] One embodiment of the automotive transmission system is
shown diagrammatically in FIG. 65. It consists of: [0170] 1) An
underdrive CVT to perform the normal transmission ratio changing an
internal combustion powered vehicle needs. [0171] 2) An overdrive
CVT to translate vehicle kinetic energy into the acceleration of a
flywheel. [0172] 3) A flywheel to store vehicle kinetic energy.
[0173] 4) Three clutches to change operating modes of the
system.
[0174] The operating modes of this system are summarized in Table
1. Since in this system the engine is too small to accelerate the
vehicle as fast as most customers expect, the first step necessary
before starting a trip is for the engine to spin up the flywheel
while the vehicle is stationary.
TABLE-US-00001 TABLE 1 Underdrive Overdrive Driving Mode Clutch A
Clutch B Clutch C CVT ratio CVT ratio Engine revs flywheel, vehicle
X 1:1 1:1, then increases to stationary 0:1 Vehicle accelerates
with flywheel X X 1:0, then increases to 1:1 1:1 Vehicle
accelerates with engine X X 1:0, then increases to 1:1 1:1 Vehicle
decelerates with flywheel X NA 1:1, then increases to 0:1 Vehicle
decelerates with engine X X X 1:0 1:1, then increases slightly
Flywheel revs during cruise X X Speed dependant 1:1, then increases
to 0:1 Engine reverses vehicle X X X 1:0 reverses "X" signifies
that the indicated clutch is engaged
[0175] The engine is connected to the input of the underdrive
transmission through clutch A. The underdrive transmission remains
at a 1:1 ratio. Its output is geared to the input of the overdrive
transmission, which slowly increases its overdrive ratio from 1:1
up to 1:100 and beyond. This accelerates the flywheel to high
speed.
[0176] The next mode involves using flywheel energy to accelerate
the vehicle to cruising speed. Here clutch B is engaged, connecting
the flywheels high rotational velocity to the underdrive
transmission which is initially set to the 1:0 ratio. The output
does not turn regardless of how fast the input turns. Clutch C is
also engaged to transmit the transmissions output speed to the
vehicle drive axle. To accelerate, the underdrive ratio is now
slewed from 1:0 toward 1:1. The flywheel will decelerate rapidly
while the vehicle accelerates.
[0177] Once the ratio is at 1:1 the flywheel has transferred all
the energy it can to the vehicle. Depending on the speed desired by
the driver of the vehicle, the acceleration would typically cease
long before the flywheel is exhausted. The engine is now ready to
maintain the cruise condition.
[0178] For the engine to maintain cruise, clutch A and C are
engaged. The engine drives the vehicle through the underdrive
transmission, which is set at the appropriate ratio for the speed
desired. Modest accelerations are possible with the small engine
through typical control of the throttle and transmission ratio.
[0179] When it comes time to decelerate the vehicle, clutch A is
released and the overdrive transmission starts to slew its ratio
up--accelerating the flywheel. Higher vehicle deceleration rates
can be obtained by increasing the overdrive ratio at a faster
rate.
[0180] If a very long mountain grade is encountered that requires
engine braking, all three clutches are engaged and the underdrive
transmission is set to 1:0 ratio--putting it into neutral. The
overdrive CVT is initially set to the 1:1 ratio. At this point the
vehicle is back driving through the overdrive CVT to the engine.
More engine braking can be obtained by increasing the ratio of the
overdrive CVT. This can be used to increase engine braking up to
the maximum speed of the engine.
[0181] There might be times when it is desirable to spin up the
flywheel while the engine is driving the vehicle at cruising speed.
This would be done if performance was more a requirement than
economy and the driver always wanted the vehicle ready for another
burst of acceleration (Performance Mode Driving). This would be
accomplished with the same clutch A and C combination as normal
cruising, but now the overdrive CVT slowly ramps up its ratio to
accelerate the flywheel, but not so fast as to overpower the small
engine and slow down the vehicle.
[0182] For vehicle reversing, the same clutch state of A, B and C
engaged and underdrive CVT in neutral (1:0) ratio as the engine
braking case is set. The one difference is the overdrive CVT is
over ratio-ed into reverse mode as described earlier. The input and
output invert when doing so, but that is just what we want as we
are now driving backwards from the engine to the wheels through the
overdrive CVT.
[0183] Single-CVT Full Automotive Transmission System
[0184] FIG. 66 and Table 2 illustrate an alternate embodiment to
the above system that only requires one CVT unit that is switchable
from overdrive to underdrive mode as described earlier. This system
requires an output reversing gear unit but overall saves
considerable hardware over the dual CVT system. The one
disadvantage with this system is that some finite time is required
to shift the unit from overdrive to underdrive mode. This could
lead to some lost opportunity to capture braking energy if the
driver switches back and forth from accelerating to braking too
rapidly. The first system described with separate overdrive and
underdrive CVT's can always have the ratios set at the right point
to capture or release stored kinetic energy on an instants
notice.
[0185] Referring to FIG. 66 and Table 2, one can see that all
driving modes of the more "deluxe" version 1 system described above
are obtainable.
[0186] The engine can be used to pre-accelerate the flywheel while
the vehicle is stationary by applying clutch B and setting the
transmission to it the 1:1 ratio in overdrive mode. The engine can
then back drive through the CVT to rev up the flywheel.
[0187] Next the flywheel's kinetic energy can be used accelerate
the vehicle by setting the transmission to the 1:0 ratio in
underdrive mode and engaging band B. Band B engages the output
reverser to negate the rotation reversal of the CVT in Underdrive
mode. The CVT's ratio can now be slewed from 1:0 up to 1:1 to
accelerate the vehicle and decelerate the flywheel.
[0188] Once the flywheel's kinetic energy is nearly exhausted, the
clutch A can be engaged to use the remaining kinetic energy in the
flywheel to start the engine. Once running, the engine can maintain
the speed of the vehicle or perform modest accelerations through
clutch A, the CVT and the output reverser.
TABLE-US-00002 TABLE 2 Under/Over Driving Mode Clutch A Clutch B
Clutch C Band E mode CVT ratio Engine revs flywheel, vehicle X O
1:1, then increases to stationary 0:1 Vehicle accelerates with
flywheel X U 1:0, then increases to 1:1 Vehicle accelerates with
engine X X U 1:0, then increases to 1:1 Vehicle decelerates with
flywheel X O 1:1, then increases to 0:1 Vehicle decelerates with
engine X X O 1:1, then increases slightly Flywheel revs during
cruise X X O 1:1, then increases to 0:1 Engine reverses vehicle X X
U 1:0, then increases slightly
[0189] If the vehicle is moving at speed and a deceleration is
called for, the vehicle's linear kinetic energy can be fully
captured by transferring it into accelerating the flywheel. This is
accomplished by setting the CVT ratio to 1:1 in overdrive mode and
engaging clutch C. The drive wheels can now back drive the CVT and
as the ratio is slewed up to 0:1, the flywheel will accelerate and
the vehicle will decelerate. If the vehicle needs to decelerate for
a longer period of time than the flywheel can safely absorb without
over speeding, either clutch A or clutch C can be engaged to
utilize engine braking to retard the vehicle without a time
limitation. Clutch B would be engaged if the flywheel is already
absorbed maximum energy and is spinning at a speed beyond the
operating speed of the engine. If the vehicle were previously set
to "Hilly Driving" mode, the control system would engage clutch A
immediately upon decelerating before the flywheel reached to high
of a speed. This way, the CVT ratio could be used to over run the
engine at a higher speed than the fixed ratio connection of clutch
B for more effective engine braking without the fixed energy
absorption limit of the flywheel alone.
[0190] For purposes of "Performance Mode" driving, the flywheel can
be revved to maximum speed while the vehicle is cruising under
engine power to be ready at all times for a burst of high
acceleration. This accomplished by driving the vehicle with the
engine through the direct connection of clutch B and the output
reverser set to a direct drive with clutch C, while the CVT, in
overdrive mode, accelerates the flywheel.
[0191] Reverse can be accomplished with either the engine or
flywheel by using the same settings as forward acceleration but
with the opposite setting of the output reversing gearbox.
EXAMPLE
[0192] A prototype transmission was constructed and mounted on a
test rig for the purpose of measuring power transmission
efficiency. The transmission has eight output elements as
diagrammed in FIG. 9 and having a gear train as shown in FIG. 46
was constructed and tested for efficiency. A one horsepower
servomotor drove the input 10 of the transmission and an identical
servomotor absorbed the output power. Speed on the input and output
was measured directly from the servomotor encoders and the torque
was assumed to be proportional to the voltage delivered to each
servomotor. Torque (voltage) was commanded on a 0 to 7
non-dimensional scale. By calculation, the maximum torque (7)
should be about 5 ft-lbs. A typical efficiency test was conducted
by setting the torque on the output motor to a fixed resistance
level of 3, 5 or 7. The input motor was commanded to run to a fixed
speed of 1000 rpm. Various transmission speed ratios and output
torques were measured. During the test, we measured the resulting
torque delivered by the servo controller to maintain 1000 rpm on
the input and measure the resulting speed from the output encoder.
By multiplying speed times torque, input and output power can be
calculated and divided to yield power transmission efficiency.
Small torque offsets were subtracted from measured torques in the
calculation to account for the servo motors internal spin losses.
These were measured by spinning the motors alone without the
transmission in place.
[0193] The data is graphed in FIG. 67. There is fairly large data
scatter as this is a rudimentary way to measure rotating torque.
Data is shown for the three output torque levels (3, 5, and 7) and
speed ratios from approximately 1:1 to 40:1. Linear regression
lines are calculated and displayed to attempt to look through the
data scatter. As expected, the transmission is most efficient at
higher torque levels and lower speed ratios. But even at a 40:1
reduction ratio, the power transmission efficiency is
encouraging.
LIST OF REFERENCE NUMBERS
[0194] 10 Input shaft 130 Second output sprocket [0195] 20 Input
disk 140 Chain or belt [0196] 30 Crank pin 142 Final drive shaft
[0197] 40 Connecting rod 143 Reversing gearbox [0198] 50 Bell crank
145 Sliding crank pin plate [0199] 60 One-way clutch 150 Rotatably
movable disk [0200] 70 Output shaft 155 Protruding dog element
[0201] 80 Slot in scotch yoke 160 Piston [0202] 90 Scotch yoke 170
Cylinder [0203] 93 Secondary link 180 Adjustment sleeve [0204] 94
First secondary bell crank 190 Case ground [0205] 95 Second
secondary bell crank 200 External adjustment mechanism [0206] 100
Main backing plate 210 Roller bearing [0207] 110 Driveshaft 220
Overrun/reversing unit [0208] 120 First output sprocket 230
Electric motor
[0209] While the foregoing invention has been described in some
detail for purposes of clarity and understanding, it will be clear
to one skilled in the art from a reading of this disclosure that
various changes in form and detail can be made without departing
from the true scope of the invention. For example, all the
techniques and apparatus described above may be used in various
combinations. All publications, patents, patent applications, or
other documents cited in this application are incorporated by
reference in their entirety for all purposes to the same extent as
if each individual publication, patent, patent application, or
other document were individually indicated to be incorporated by
reference for all purposes.
* * * * *